Such rotary valves typically include an input-shaft which incorporates in its outer periphery a plurality of blind-ended, axially extending grooves separated by lands. Journalled on the input-shaft is a sleeve having in its bore an array of axially extending blind-ended slots circumferentially aligned with the lands on the input-shaft. The interfaces between the coacting input-shaft grooves and sleeve slots define axially extending orifices which open and close when relative rotation occurs between the input-shaft and the sleeve. The sides of the input-shaft grooves are contoured so as to provide a specific orifice configuration and are referred to as metering edge contours. These orifices are ported as a network such that they form sets of hydraulic Wheatstone bridges which act in parallel. Such hydraulic Wheatstone bridges are analogous in operation to conventional electrical Wheatstone bridges.
Drilled passages in the input-shaft and sleeve, together with circumferential grooves in the periphery of the sleeve, serve to communicate oil between the grooves in the input-shaft and the slots in the sleeve, a hydraulic pump, and right-hand and left-hand hydraulic assist cylinder chambers incorporated in the steering gear.
A torsion bar incorporated in the input-shaft serves to urge the input-shaft and sleeve towards a neutral, centred position when no power assistance is required. When input torque is applied by the driver to the steering wheel, the torsion bar deflects, causing relative rotation of the sleeve and input-shaft from the neutral position. This so called "valve operating angle" imbalances the sets of hydraulic Wheatstone bridges and hence causes a differential pressure to be developed between the right-hand and left-hand cylinder chambers. The "boost characteristic" of the rotary valve, that is the functional relationship between the above mentioned input torque and differential pressure, is largely determined for a given steering gear application by the geometry of the metering edge contours.
Traditionally the networl(of orifices in a rotary valve employ 2, 3 or 4 Wheatstone bridges, necessitating respectively 4, 6 or 8 input-shaft grooves and sleeve slots. Each Wheatstone bridge comprises a right-hand and a left-hand oil flow path, henceforth termed "limbs", and each right-hand and left-hand limb in turn comprises upper and lower portions. The upper and lower portions of each right-hand and left-hand limb meet respectively at a point of connection to the right-hand and left-hand cylinder chamber, henceforth termed the right-hand and left-hand "cylinder ports" of the valve.
In the neutral position of the rotary valve, oil from the hydraulic pump divides and enters each Wheatstone bridge at the valve "inlet port". At this point flow further divides and enters the upper right-hand and left-hand limbs, each containing an "inlet orifice". After being metered through such inlet orifices, oil is communicated to the respective cylinder ports and to the respective interconnection to the lower limbs. Depending on the intercylinder flow rate drawn by the motion of the piston in the cylinder chamber, oil continues to flow through the tower right-hand and left-hand limbs, metering through a "return orifice" in each limb, and recombining immediately upstream of the "return port" of the rotary valve.
The networl of two inlet orifices and two return orifices, constituting each Wheatstone bridge, is ported in the rotary valve such that, for a given relative angular displacement of the input-shaft and sleeve from their neutral position, mutually opposite orifices on each Wheatstone bridge simultaneously close or open. For example, the left-hand inlet and right-hand return orifices both close (ie. increase in restriction to oil flow) while the right-hand inlet and left-hand return orifices both open (ie. decrease in restriction to oil flow). According to classical Wheatstone bridge theory, for a given oil flow through each Wheatstone bridge, a differential pressure is therefore developed between the right-hand and left-hand cylinder ports, providing the necessary level of power assistance for each value of valve operating angle.
The general method of operation of such traditional rotary valves is well known in the prior art of power steering design and described in greater detail in U.S. Pat. No. 3,022,772 (Zeigler et al), commonly held as being the "original" patent describing the rotary valve concept. Rotary valves of this format will henceforth be termed "direct mode valves" since all Wheatstone bridges within the valve incorporate direct hydraulic communication to the cylinder ports.
Rotary valves are nowadays regularly incorporated in firewall-mounted rack and pinion steering gears, and in this situation, any noises such as hiss emanating from the valve are very apparent to the driver. Hiss results from cavitation of the hydraulic oil as it flows in the orifices defined by the metering edge contours and the adjacent edges of the sleeve slots, particularly during times of high pressure operation of the valve such as during parking, where differential pressures of 8-10 MPa or more can be generated. It is well known in the art of power steering valves that an orifice is less prone to cavitation if the associated metering edge contour has a high aspect ratio of axial length to radial depth, thereby constraining the oil to flow as a thin sheet of constant depth along the full axial extent of one metering edge contour and if, furthermore, the flow of oil is evenly divided amongst several metering edge contours ported to act in parallel, so further effectively reducing the flow of oil that may flow through any one orifice. It is also well known that cavitation is less likely to occur if the metering edge contour, where it intersects the outside diameter of the input-shaft, is nearly tangential thereto, hence constituting a shallow chamfer typically inclined at an angle of between 4 deg and 8 deg.
Such shallow chamfers have been widely used in rotary valves for noise suppression over the last 20 years. In order to achieve the necessary depth and form accuracy, these chamfers are normally ground in special indexable or cam- type grinding machines resulting in long overall cycle times, relatively expensive capital equipment, and hence high overall manufacturing cost.
Another requirement which is increasingly becoming accepted for the design of rotary valves is the need for a linear boost characteristic. During vehicle cornering, it is advantageous that a substantially linear relationship exists between the driver's input torque and the differential pressure associated with such a cornering manoeuvre. This leads to the sensation of "progression" in the power assistance and maximises steering feel in such critical situations. Associated with the requirement for a linear boost characteristic, it is also highly desirable to maximise the extent of the linear region before the maximum parking torque (and hence parking pressure) is reached. This necessitates a fast transition or "turn-around" of the linear boost characteristic to a region of much steeper slope associated with the higher differential pressures used for parking. For a given slope and extent of the linear boost characteristic of the rotary valve during cornering, the torque required to be exerted by the driver during parking is therefore minimised.
Chamfer type metering edge contours can, to some limited degree, generate a low noise linear boost characteristic if the chamfer is designed as a scroll, as disclosed in U.S. Pat. No. 5,267,588 (Bishop et al), or as a series of flat facets, as disclosed in U.S. Pat. No. 4,460,016 (Haga et al). However, in both these cases, the extent of the linear boost region is relatively short and the transition to the steeper parking region of the boost characteristic is prolonged, and therefore not optimal in terms of minimising parking torque.
Another technique, well known in the art, for suppressing valve noise in power steering valves is the application of back pressure to an otherwise cavitating orifice, thereby raising pressures within the orifice above the vapour pressure of the hydraulic oil and hence preventing the onset of cavitation. Chamfer style metering edge contours need not necessarily be used on the input-shaft if this alternative method of noise suppression is employed. Much steeper and axially shorter metering edge contours can in fact be used, contours which would otherwise be excessively noisy in the absence of such back pressure. Such steeper and generally more complex shaped metering edge contours can be manufactured by coining, roll-imprinting or traditional hobbing methods and, if appropriately designed, can generate the previously described desirable linear boost characteristic with a fast turn-around. U.S. Pat. No. 4,335,749 (Walter) shows a direct mode valve incorporating an extra (second) orifice in the lower portions of the left-hand and right-hand limbs of each bridge. This orifice progressively closes with increasing valve operating angle until a constant orifice area is reached which, based on the flow through the limb, applies a predetermined back pressure to the closing upstream inlet orifice. Such a valve format is based on 6 orifices per bridge and, if 3 bridges are employed in the valve, requires 9 input-shaft grooves and 9 sleeve slots. If 4 bridges are employed (as in the case of a traditional 8 groove/slot rotary valve), 12 input-shaft grooves and 12 sleeve slots are required. This format is therefore non-standard and requires extra manufacturing cost.
Further however, experiments have shown that elimination of cavitation noise in an orifice generating 10 MPa (say) differential pressure requires a downstream back pressure to be applied which is as much as 1 MPa or more . If such high levels of back pressure were generated by the return orifices according to the methodology disclosed in U.S. Pat. No. 4,335,749 (Walter), this back pressure would raise the inlet pressure required to be supplied by the hydraulic pump by the same 1 MPa, without any of this additional pressure being applied differentially at the cylinder chamber. This is because this direct mode valve arrangement contains cylinder port connections in every bridge and the return orifice used for back pressure generation is downstream of such connections. The 1 MPa increase in valve inlet pressure would be totally wasted in terms of generating power assistance force and would simply raise the operating pressure of the hydraulic pump. The latter situation is highly undesirable since energy loss in the hydraulic system is proportionally increased. Also pump noise, leakage and potential hydraulic line failure all become bigger problems as the pump relief valve setting is necessarily increased to accommodate the increased valve operating pressure, for example from 10 MPa to 11 MPa in this case.
For this reason the practical level of back pressure that can be applied by the return orifice according to the above prior art invention is limited to about 300-400 kPa, well short of the 1 MPa or more needed to substantially eliminate cavitation noise through the operating pressure range of the power steering valve.
Another class of rotary valve, henceforth termed "bypass mode valves", is quite distinct from the class of direct mode valves earlier described. Bypass mode valves also utilise parallel arrangements of Wheatstone bridges, however not all bridges in this case contain a hydraulic connection to a cylinder port between the inlet and return orifices. The bridges which employ a cylinder connection will henceforth be termed "primary bridges" and those which don't employ a cylinder connection termed "secondary bridges". In the latter case the left-hand and right-hand limbs contain one or more inlet and return orifices but with no interposed cylinder port connection. In this manner, for certain valve operating angles, hydraulic oil at least partially by-passes the primary bridge(s) which incorporate the connection to the cylinder.
Such bypass mode valves were first put forward for speed sensitive power steering applications. For example, arrangements described in U.S. Pat. Nos. 4,570,735 (Duffy) and 4,570,736 (Waldorf) and Japanese Patent 04-031175 (Suzuki et al) involve a bypass mode valve with an electronically modulated variable orifice residing in the inlet to the secondary bridges and modulated as a function of vehicle speed. Other later arrangements such as shown in Japanese Patent 02-306878 (Suzuki) and U.S. Pat. No. 5,092,418 (Suzuki et al) use an electronically modulated variable orifice residing in the return line from the secondary bridges. In such speed sensitive applications, the degree of bypass of hydraulic oil through the secondary bridges is used to control the boost characteristic as a function of vehicle speed.
Bypass mode valves have also been utilised in a non-speed sensitive format to improve the linearity and produce a fast turn-around of the boost characteristic for valves employing chamfered metering edge contours. For example Japanese Patents 04-031176, 05-042880 and 06-278623 (all Suzuki et al) and U.S. Pat. No. 4,470,432 (Kervagoret) show orifice networks very similar to the abovementioned speed sensitive applications except that the electronically modulated variable orifice is now a fixed "drill-hole" style orifice either upstream or downstream of the secondary bridges. In situations where relatively conventional metering edges are used in the orifices of the secondary bridges, such arrangements will tend to be noisy for two reasons. Firstly the very low aspect ratio of the fixed orifices (ie. unity for a drill hole will be a source of cavitation for the relatively high oil flows involved. Secondly in these arrangements, for high valve operating pressures, all pump flow is communicated to the return port via only two stages of pressure drop: the restrictions offered by relevant closing secondary orifice and the fixed orifice (or vice versa).
U.S. Pat. No. 4,577,660 (Haga) shows an 8 slot by-pass valve again intended to produce a linear boost characteristic with a fast turn around. In this case the secondary inlet orifices are overlapped and in fact closed on-centre, their sudden opening off-centre intended to produce the required discontinuity in the boost characteristic. However such an arrangement, with a substantial portion of the valve closed on-centre, would naturally exhibit higher than usual back pressure on-centre and would therefore be energy inefficient.
Japanese Patent 04-292265 (Suzuki et al) shows a relatively complex by-pass mode valve employing 10 input-shaft grooves and corresponding 10 sleeve slots. An extra orifice is positioned in the lower portion of each secondary bridge limb and, as it closes, provides a similar flow redistribution function to the earlier mentioned fixed orifice. Such a valve arrangement is expensive because of the larger quantity of input-shaft grooves and sleeve slots, and the associated interconnecting porting (eg. drill holes) to hydraulically communicate such slots/grooves. Moreover 10 input-shaft grooves and 10 sleeve slots are difficult to package using a standard input-shaft outside diameter (or corresponding sleeve inside diameter), typically in the range 19.0-22.5 mm, and yet still retain sufficient inter and intra slot/groove spacing to accommodate such interconnecting porting.
However the nature of the diversion of oil flow in bypass mode valves between the primary and secondary bridges means that, when such valves generate a large differential pressure at the cylinder, essentially only the secondary bridges transmit any oil flow. This means that the individual orifices in the secondary bridges tend to be prone to cavitation noise even if shallow chamfers are employed as the metering edge contours according to the prior art. Japanese Patent 05-310136 (Suzuki et al) proposes to reduce this problem by employing an electronically modulated variable orifice positioned at the return port of a bypass mode valve, this variable orifice controlled to produce a restriction (and hence generate back pressure) as a function of the sensed inlet pressure to the valve. For reasons earlier described, such an arrangement is energy inefficient and, moreover in this case, adds significant cost to the power steering system.
Nevertheless, bypass mode valves do offer a major advantage over direct mode valves in that, any back pressure applied within the secondary bridge network to suppress cavitation noise generated by the respective orifices does not raise the overall inlet pressure to the valve, as supplied by the hydraulic pump, for a given differential pressure applied at the cylinder. Hence such back pressure is not wasteful in terms of energy and in fact is usefully used to generate some portion of the power assistance at the cylinder. There is therefore no need to use higher pump relief valve settings and the previously referred to large levels of back pressure (eg. 1 MPa) can be theoretically utilised to substantially eliminate cavitation noise without any major disadvantage in terms of valve function.
The first and second aspects of the present invention are directed at utilising the above mentioned benefits of bypass mode valves, and yet provide low levels of cavitation noise in a rotary valve without necessarily increasing the number of input-shaft grooves or sleeve slots. Another aim is to permit such low levels of cavitation noise to be achieved using metering edge contours of the more steeply sloping variety earlier referred to. Such metering edge contours can be produced by coining, roll-imprinting or hobbing and not only offer significant cost savings compared to shallow chamfers which generally must be ground, but also enable much more flexibility in the design of the boost characteristic, particularly the provision of a linear boost characteristic with a fast turn-around. Also such steeply sloping metering edge contours can generally be designed to be axially shorter than the comparable shallow chamfer, enabling the overall rotary valve package to be likewise shortened.
The first aspect of the present invention consists in a rotary valve for a hydraulic power steering gear comprising a valve housing having an inlet port to receive hydraulic fluid from a pump, a return port to return hydraulic fluid to the pump, and cylinder ports to communicate hydraulic fluid to left and right-hand cylinder chambers of the power steering gear, the valve also comprising an input-shaft having in its outer periphery a plurality of axially extending grooves separated by lands, a sleeve journalled on said input-shaft, said sleeve having in its bore an array of axially extending slots circumferentially aligned with the lands on the input-shaft, the interfaces between the coacting input-shaft grooves and sleeve slots defining axially extending orifices controlling fluid flow within the valve, the orifices opening and closing when relative rotation occurs between the input-shaft and sleeve from a neutral position, the orifices being ported as a network such that they form one or more primary and one or more secondary hydraulic Wheatstone bridges arranged in parallel, each said bridge comprising two limbs hydraulically communicating the inlet and return ports, each said limb containing an inlet orifice hydraulically communicating to the inlet port and a return orifice hydraulically communicating to the return port, the magnitude of the hydraulic flow through each bridge varying in accordance with the restriction offered by the respective inlet and return orifices in that bridge, the limbs of the primary bridge incorporating means providing hydraulic communication to one of the cylinder ports at a point of interconnection of the respective inlet and return orifices in that limb, the limbs of the secondary bridge not incorporating means providing hydraulic communication to the cylinder ports, characterised in that the return orifice in each limb of said secondary bridge is formed by a metering edge contour on the edge of the secondary return groove associated with said return orifice, said metering edge contour circumferentially overlapping the adjacent sleeve bore land when the rotary valve is in its neutral position to such an extent that said return orifice provides a restriction to hydraulic flow as the upstream inlet orifice in the same limb closes for all valve operating angles from said neutral position, said return orifice applying a back pressure to said upstream inlet orifice, said back pressure being sufficient to significantly suppress the generation of cavitation noise in said inlet orifice.
It is preferred that a substantially constant restriction area is provided by said return orifice as the upstream inlet orifice in the same limb closes for all valve operating angles from said neutral position.
It is preferred that the input-shaft metering edge contour employed in said return orifice is formed in cross-section such that a region of locally reduced metering edge depth lies in the overlapped region of the coacting input-shaft metering edge contour and adjacent sleeve bore land, that is in the region lying radially inside the adjacent sleeve land. Said substantially constant restriction area provided by said return orifice can be considered as constituting a hydraulic throat which serves to significantly suppress cavitation noise or turbulence as the hydraulic oil flows past the adjacent sleeve edge and enters this return orifice.
It is preferred that cavitation and other flow noise can be further reduced by raising the back pressure downstream of the secondary return orifice.
It is preferred that hydraulic flow from the primary bridge is hydraulically communicated to the return port via a primary return path and the hydraulic flow from the secondary bridge is hydraulically communicated to the return port via a secondary return path, a restriction existing in the secondary return path.
In first and second embodiments it is preferred that the secondary return path passes through the bore of the input-shaft.
In a first embodiment it is preferred that the radial holes which hydraulically communicate the secondary return groove to the input-shaft bore are reduced in diameter, thereby generating back pressure in the secondary return groove downstream of the secondary return orifice.
The capability of these radial holes to apply such back pressure without themselves causing a noise problem can be further enhanced if conical or tapered entries are employed on these holes. This can be readily and cheaply achieved via a step form on the drill used to machine these holes or by laser erosion.
In the case of this first embodiment, it is also preferred that the primary return path also passes through the bore of the input-shaft. However the radial holes which hydraulically communicate the primary return grooves to the input-shaft bore are sufficiently large in diameter in this case that no substantial restriction is generated.
In a second embodiment, and also in a later referred to fifth embodiment, the restriction existing in the secondary return path is preferably annular in geometry. It is preferred that this annular restriction has a cross-section to flow which has a high aspect ratio, in order to suppress its generation of cavitation noise. In this second embodiment it is preferred that only the secondary return path passes through the bore of the input-shaft. Oil entering the input-shaft bore via the aforementioned radial holes is restricted using a diametrically enlarged portion on the torsion bar. This enlarged portion is arranged to have a small radial clearance with respect to the input-shaft bore, hence creating an annular restriction for hydraulic oil as it flows axially in this bore towards the return port of the valve housing.
According to this embodiment, hydraulic oil from the primary bridge is ported directly to the return port so that it is not required to flow through the input-shaft bore, and hence is not subject to this additional restriction. This is achieved by axially extending the input-shaft grooves associated with the primary return orifices in the form of channels, allowing hydraulic oil flow in the primary bridge to exit directly axially from these grooves through these channels.
The diametrically enlarged portion of the torsion bar can be integrally machined as part of the torsion bar during its manufacture. However, in order to maximise the working length of the reduced diameter portion of the torsion bar, and hence lower the maximum stress endured by the torsion bar for a given working diameter and torsional spring rate, the diametrically enlarged portion on the torsion bar is preferably formed as an annular bush which is plastic moulded around the metallic portion of the torsion bar as a separate subsequent operation. The plastic material must be chemically resistant to hydraulic oil and is preferably an engineering plastic such as Delrin.RTM. or Lurathane.RTM..
If the annular bush is made to additionally extend axially such that it overlaps the secondary return radial holes in the input-shaft, the use of such compliant plastic material for this bush has been found to assist dampening the hydraulic turbulence noise generated by the secondary return oil as it flows radially into the input-shaft bore and thence necessarily turns perpendicularly to continue flowing axially down this bore.
In a third embodiment it is preferred that the secondary return path does not pass through the bore of the input-shaft. Axially extending secondary return channels are formed in the sleeve bore which are circumferentially aligned with the secondary return grooves. The channels extend to the axial extremity of the sleeve bore and arranged to communicate hydraulic fluid to the return port. The radial depth of the channels is small, thereby interacting with the adjacent outside diameter of the input-shaft to form a high aspect ratio restriction in the secondary return path downstream of the secondary return grooves. It is preferred that at least one secondary return channel extends to both axial extremities of the sleeve bore.
It is also preferred that axially extending primary return channels are formed in the sleeve bore and arranged to be circumferentially aligned with the primary return grooves. These additional channels also extend to the axial extremity of the sleeve bore and are arranged to communicate hydraulic fluid to the return port. It is preferred that the radial depth of the primary return channels is larger than that of the aforementioned shallow secondary return channels since no restriction is required to be generated in the primary return path. It is also preferred that at least one of the primary return channels extends to both axial extremities of the sleeve bore.
For reasons of ease of manufacture, it is preferred that all primary and secondary return channels extend to both axial extremities of the sleeve bore, enabling all such channels to be formed with a single multi-tooth broaching tool.
In a fourth embodiment it is also preferred that the secondary return path does not pass through the bore of the input-shaft. The secondary return grooves are axially extended as shallow, high aspect ratio channels formed via their interaction with the adjacent sleeve bore. These channels extend to the axial extremity of the sleeve bore, thereby providing a restriction in the secondary return path. It is also preferred that the primary return grooves are similarly axially extended as radially deeper channels to facilitate a relatively unrestricted primary return path.
In a fifth embodiment is is also preferred that the secondary return path does not pass through the bore of the input-shaft. The secondary return grooves are axially extended in at least one direction to communicate with an annular cavity formed by the interaction of a reduced diameter portion of the input-shaft outer periphery and the sleeve bore. The annular cavity acts as a manifold to gather secondary return oil flow, which is then communicated via an annular restriction to the return port. The annular restriction is preferably generated by a predetermined small radial clearance existing between the above mentioned reduced diameter portion of the input-shaft and the inside diameter of a radially inwardly extending portion of the sleeve bore. Preferably the radially inwardly extending portion of the sleeve bore is formed as an accurately internally and externally sized annular pressed-metal cup which is press-fitted inside the sleeve skirt to seal against the axial extremity of the sleeve bore. Preferably the predetermined radial clearance is such that the resulting annular restriction has a high aspect ratio in order to suppress its generation of cavitation noise. It is also preferred that the primary return path passes through the bore of the input-shaft in a similar manner to that described in reference to the first embodiment, thereby bypassing the annular restriction en-route to the return port.
It is preferred that the rotary valve has eight input-shaft grooves.
It is preferred that the rotary valve has eight sleeve slots.
The second aspect of the present invention consists in a rotary valve for a hydraulic power steering gear comprising a valve housing having an inlet port to receive hydraulic fluid from a pump, a return port to return hydraulic fluid to the pump, and cylinder ports to communicate hydraulic fluid to left and right-hand cylinder chambers of the power steering gear, the valve also comprising an input-shaft having in its outer periphery a plurality of axially extending grooves separated by lands, a sleeve journalled on said input-shaft and rotationally secured to a driven member, said sleeve having in its bore an array of axially extending slots circumferentially aligned with the lands on the input-shaft, the interfaces between the coacting input-shaft grooves and sleeve slots defining axially extending orifices controlling fluid flow within the valve, the orifices opening and closing when relative rotation occurs between the input-shaft and sleeve from a neutral position, a torsion bar residing in a bore of the input-shaft compliantly connecting the input-shaft and driven member, and arranged to urge the sleeve and input-shaft to the neutral position, the orifices being ported as a network such that they form one or more primary and one or more secondary hydraulic Wheatstone bridges arranged in parallel, each said bridge comprising two limbs hydraulically communicating the inlet and return ports, each said limb containing an inlet orifice hydraulically communicating to the inlet port and a return orifice hydraulically communicating to the return port, the magnitude of the hydraulic flow through each bridge varying in accordance with the restriction offered by the respective inlet and return orifices in that bridge, the limbs of the primary bridge incorporating means providing hydraulic communication to one of the cylinder ports at a point of interconnection of the respective inlet and return orifices in that limb, the limbs of the secondary bridge not incorporating means providing hydraulic communication to the cylinder ports, characterised in that hydraulic flow from said primary bridge is hydraulically communicated to the return port via a primary return path and the hydraulic flow from the secondary bridge is hydraulically communicated to the return port via a secondary return path, an annular restriction existing in the secondary return path.
It is preferred that the annular restriction existing in the secondary return path has a cross-section to flow which has a high aspect ratio.
It is preferred that the aspect ratio be greater than 10.
It is preferred that one but not both of the primary or secondary return paths passes through the bore of the input-shaft.
In a first embodiment it is preferred that the secondary return path passes through the bore of the input-shaft and the annular restriction is formed within this bore.
It is preferred that the annular restriction formed in the bore of the input-shaft is generated by virtue of a small radial clearance existing between a diametrically enlarged portion of the torsion bar and the input-shaft bore.
It is preferred that hydraulic flow from the primary bridge is directly communicated to the return port via channels formed as an axial extension of the input-shaft grooves associated with the primary return orifices. Because this hydraulic flow is not communicated through the input-shaft bore, it is not subject to the abovementioned annular restriction.
Various preferred embodiments are possible for the geometry and construction of the diametrically enlarged portion of the torsion bar and have already been described in reference to the first aspect of the present invention.
In a second embodiment is is preferred that the secondary return path does not pass through the bore of the input-shaft and the annular restriction is formed at the input-shaft/sleeve interface. The secondary return grooves are axially extended in at least one direction to communicate with an annular cavity formed by the interaction of a reduced-diameter portion of the input-shaft outer periphery and the sleeve bore. The annular cavity acts as a manifold to gather secondary return oil flow, which is then communicated via an annular restriction to the return port. The annular restriction is preferably generated by a predetermined small radial clearance existing between the above mentioned reduced diameter portion of the input-shaft and the inside diameter of a radially inwardly extending portion of the sleeve bore. Preferably the radially inwardly extending portion of the sleeve bore is formed as an accurately internally and externally sized annular pressed metal cup which press-fitted inside the sleeve skirt to seal against the axial extremity of the sleeve bore. Preferably the predetermined radial clearance is such that the resulting annular restriction has a high aspect ratio in order to suppress its generation of cavitation noise. It is also preferred that the primary return path passes through the bore of the input-shaft in a similar manner to that described in reference to the first embodiment of the first aspect of the present invention, thereby bypassing the annular restriction en-route to the return port.
In the case of both first and second embodiments of the second aspect of the present invention, hydraulic flow from the secondary bridge passes axially through the relevant annular restrictions and hence applies a back pressure to all secondary orifices upstream of this restriction. The restriction area is therefore substantially constant and arranged to provide sufficient back pressure to suppress the generation of cavitation noise in these secondary orifices for all valve operating angles.
It is preferred that the rotary valve has eight input-shaft grooves.
It is preferred that the rotary valve has eight sleeve slots.